Cryogenic refrigeration of a process medium

ABSTRACT

The present invention pertains to a cryogenic refrigeration system and method for cryogenic refrigeration of a process medium. In particular, the invention relates to a counter flow heat exchanger configuration and pressure regulator arrangement to reduce exergetic losses in the system. Accordingly, a cryogenic refrigeration system is suggested comprising a conduit (2) configured to provide a supply flow (10) of a process medium, a counter flow heat exchanger (3), which is thermally coupled to a heat exchanger section (2A) of the conduit (2) and comprises an inlet (34) at a cold end (30) of the heat exchanger (3) and an outlet (36) at the warm end (32) of the heat exchanger (3), a first pressure regulator (4), which is in fluid communication with the conduit (2) and is arranged downstream of the heat exchanger section (2A), and a vessel (5), which is in fluid communication with the conduit (2) and is arranged downstream of the first pressure regulator (4), wherein the vessel (5) is in fluid communication with the inlet (34) of the heat exchanger (3) and is configured to provide an evaporated gas flow from the process medium to the inlet (34) of the heat exchanger (3). Furthermore, the conduit (2) is free of any evaporation heat exchanger upstream of the heat exchanger section (2A) of the conduit (2).

TECHNICAL FIELD

The invention relates to a system and a method for cryogenicrefrigeration of a process medium. In particular, the invention relatesto a counter flow heat exchanger configuration and pressure regulatorarrangement to reduce exergetic losses in the system.

TECHNOLOGICAL BACKGROUND

Refrigeration plants providing an isothermal load below the saturationtemperature of the process medium at atmospheric pressure are commonlyimplemented by subcooling the supply flow by means of a counter flowheat exchanger configured as an evaporator. For example, for helium, theload may be provided below 4.4 K while the supply flow is generallyprovided above atmospheric pressure. In evaporator heat exchangers apart of the liquid phase from the supply flow above atmosphericpressure, e.g. between 1.05 and 1.50 bar, is supplied to a turbine, acontrol valve or similar expansion device and enters the heat exchangerand partly evaporates, wherein the evaporated gas is released into theheat exchanger in a warmer temperature level and the liquid isrecirculated, i.e. the liquid phase exiting the evaporator heatexchanger reenters the evaporated heat exchanger at the entry thereof.Accordingly, when using helium as a process medium, a liquid phasetemperature may be provided between e.g. 4.26 and 4.67 K while thetemperature of the supply flow is between 4.3 and 4.7 K. The supply flowmay then be further cooled in another downstream counter flow heatexchanger.

Although the implementation of an evaporator heat exchanger may providea pre-cooling of the supply flow, such implementation has severaldisadvantages. For example, exergetic losses occur due to turbine andheat exchanger inefficiencies. Such exergetic losses may cause over 95%of the irreversibilities occurring in a typical helium refrigerationcoldbox. Furthermore, the refrigeration cycle comprises a largetemperature factor, e.g. 300 K and 1.0 to 4.4 K for helium, such that aneed for exergetic optimization exists to raise the efficiency of thesystem, e.g. Carnot efficiency, thereby reducing the power input to theprocess.

Furthermore, evaporator heat exchangers require recirculation of flashgas and evaporated gas on the atmospheric pressure level and furthermorerequire phase separators, both at the 4.5 K level in case of helium.Accordingly, a need exists to reduce the equipment count and sizecurrently required when using evaporator heat exchangers.

In addition, the different heat capacities of the process medium atvarious pressures below the inversion temperature cause a relativelyhigh temperature difference at the warm end of the system, i.e. betweenthe process medium exiting the system and the supply flow of the processmedium entering the system. Such temperature differences generally causeirreversibilities in the system.

SUMMARY OF THE INVENTION

It is an object of the present invention to provide an improvedcryogenic refrigeration system and a corresponding cryogenicrefrigeration method that reduces the above problems.

This object is achieved by the cryogenic refrigeration system comprisingthe features of claim 1 and the cryogenic refrigeration methodcomprising the features of claim 12. Preferred embodiments are providedin the dependent claims and by the specification and the Figures.

Accordingly, in a first aspect, a cryogenic refrigeration system issuggested, which comprises a conduit configured to provide a supply flowof a process medium and a counter flow heat exchanger, which isthermally coupled to a heat exchanger section of the conduit. The heatexchanger comprises an inlet at a cold end of the heat exchanger and anoutlet at the warm end of the heat exchanger. The system furthermorecomprises a first pressure regulator, which is in fluid communicationwith the conduit and is arranged downstream of the heat exchangersection, and a vessel, which is in fluid communication with the conduitand is arranged downstream of the first pressure regulator. The vesselis in fluid communication with the inlet of the heat exchanger and isconfigured to provide an evaporated gas flow from the process medium tothe inlet of the heat exchanger. The conduit is free of any evaporationheat exchanger upstream of the heat exchanger section of the conduit.

Accordingly, by providing a cold counter flow heat exchanger comprisingan evaporated gas flow with low specific enthalpy, the system does notrequire an evaporator to precool the supply flow. This is particularlyadvantageous when using helium, such that the system does not require anevaporating heat exchanger and phase separator at the 4.5 K level andfurthermore no recirculation of flash gas or evaporated helium on theatmospheric pressure occurs. In addition, smaller equipment such ascompressors and heat exchangers may be provided, such that thedimensions of the system may be reduced.

The cold end of the heat exchanger hence relates both to the evaporatedgas having a lower temperature and latent heat before entering the heatexchanger via the inlet and the temperature of the process medium in theconduit directly downstream of the heat exchanger section. In thiscontext the term “downstream” refers to the supply flow provided in theconduit and in relation to the initial entry of the supply flow into thesystem. Accordingly, the entry of the supply flow into the system occursupstream of the heat exchanger section. The heat exchanger section maycomprise only a part of the conduit, wherein the part of the conduitarranged upstream of the heat exchanger section and the parts of theconduit arranged downstream of the heat exchanger section and upstreamof the first pressure regulator are arranged in parallel and adjacent tothe outlet and in of the heat exchanger, respectively, to furtherimprove the heat transfer efficiency. However, the heat exchangersection and the heat exchanger may also be configured such that the heatexchanger section essentially forms the conduit, e.g. the size anddimension of the fluid couplings between the various features areminimized.

By the same token, the warm end of the heat exchanger relates to theevaporated gas that exits the heat exchanger and has absorbed heat fromthe process medium and may therefore be considered to comprise a warmertemperature and/or an increased latent heat with respect to theevaporated gas at the cold end of the heat exchanger. The warmedevaporated gas may then exit the system via the outlet at the warm endof the heat exchanger as an exhaust gas. For example, the exhaust gasmay be directly released into the atmosphere or may be retained in thesystem for further purposes and applications.

Preferably, the heat exchanger is configured to provide a temperaturefactor of the evaporated gas at the pinch point of the heat exchangerrelative to the process medium of the supply flow at the pinch point ofthe heat exchanger larger than 0.9 during normal operation of thecryogenic refrigeration system. Preferably, said temperature factor islarger than 0.98, such that temperature differences between theevaporated gas at the pinch point of the heat exchanger relative to theprocess medium of the supply flow at the pinch point of the heatexchanger are minimal and/or negligible, thereby not affecting thesystem.

Such temperature factor is possible since the system does not require anupstream evaporating heat exchanger, which generally providestemperatures of the process medium after passing the evaporating heatexchanger fixed at around e.g. 4.6K for helium, wherein the mass flowsin a steady state process at the cold and warm ends are generally equaland constant. In contrast, the cold counter flow heat exchanger mayprovide the supply flow and the evaporating gas at a higher temperaturelevel at the warm end of the heat exchanger having an increased heatcapacity, such that temperature differences may be minimized.

The above-mentioned temperature factor FT of a counter flow heatexchanger can be expressed with the temperature of the cold streamT_(c)(x) (with 0≤x≤L) and the temperature of the warm stream T_(w)(x) atthe pinch point of the heat exchanger hence where the temperaturedifference of the two streams is the smallest.

${FT} = \frac{T_{c}\left( x_{{pinch}.{point}} \right)}{T_{w}\left( x_{{pinch}.{point}} \right)}$

Alternatively, or in addition, the heat exchanger comprises an NTU(Number of Transfer Units) configured to match a temperature of theevaporated gas with a temperature of the process medium at the warm endof the heat exchanger during normal operation of the cryogenicrefrigeration system.

The implementation of a heat exchanger comprising the required NTU atleast has the advantage that the system may be thermodynamicallyoptimized while certain variables, e.g. heat exchanger parameters andboundary conditions, are not required or need not be known. Accordingly,the NTU configuration provides an alternative to the LMTD configurationto provide a thermally efficient cryogenic refrigeration system.

The term “matches” here is to be understood as essentially matching saidtemperatures and hence also includes minimal differences, e.g. up to0.05 K. For example, the area of the heat exchanger, e.g. the heattransfer area or length of the heat exchanger may be sized anddimensioned to provide the corresponding temperature range, wherein atleast the mass flow and the heat capacity values at various temperaturesof the process medium are considered to be known.

The NTU may be provided by a heat transfer area of the heat exchanger,preferably by a length of the heat exchanger. To provide a large heattransfer area the heat exchanger is preferably of a tube shape, coiledshape, and/or plate fin shape and at least partially surrounds acircumference of the conduit. This has the advantage that a largecontact surface is provided and the area may be readily increased byincreasing the length of the heat exchanger. For example, the heatexchanger may fully enclose the circumference of the conduit, wherein alongitudinal axis of the heat exchanger extending between the warm andcold end of the heat exchanger may coincide with a longitudinal axis ofthe conduit, e.g. a flowing direction of the process medium. However,instead an asymmetrical arrangement may also be provided, for example,wherein the longitudinal axis of the conduit is spaced apart from thelongitudinal axis of the heat exchanger, e.g., in a lateral arrangementto said axis. Preferably, the heat exchanger may be configured as aplate fin heat exchanger, e.g. for larger systems or plants, or as acoil finned tube heat exchanger, e.g. for smaller systems or plants.

During initial start-up of the system generally a normalization of thetemperatures and pressures is required to provide a steady state, i.e.normal operation. By matching or minimizing the temperature differenceof the process medium and the evaporated gas or exhaust gas at the warmend of the heat exchanger, the exergetic losses are reduced duringnormal operation. As such, the occurrence of irreversibilities in thesystem and the power input to the process are likewise reduced.Furthermore, by providing a cold counter flow heat exchanger having arequired NTU configuration, the system does not require an evaporator toprecool the supply flow. This is particularly advantageous when usingliquid helium, such that the system does not require an evaporating heatexchanger and phase separator at the 4.5 K level and furthermore norecirculation of flash gas or evaporated helium on the atmosphericpressure occurs. In addition, smaller equipment such as compressors andheat exchangers may be provided, such that the dimensions of the systemmay be reduced.

The increased heat transfer rate and hence cooling efficiency of theheat exchanger furthermore provides that the temperature of the supplyflow at the warm end of the heat exchanger may be much higher than thesaturation point, i.e. for helium and depending on the pressure of theprocess medium above 4.5 K and preferably as high as possible. However,said temperature range may be limited by real gas properties, such that,for e.g. helium, the temperatures are preferably between 4.5 and 20 K,more preferably between 8 and 15 K or between 10 and 13 K. Correspondinghigher temperatures may be implemented for other process media, e.g.nitrogen. Accordingly, different supply pressures above atmosphericpressure may be provided. This not only reduces the operation costs ofthe system, but also is exergetic advantageous, as heat leaked into theprocess between the main refrigeration cycle and e.g. a load occurs onan increased temperature level and hence a higher capacity of theprocess medium.

Since the pressure influences or even determines the temperature andphysical behavior of any devices or subsystems coupled to the vessel,e.g. a load such as a cryogenic user or superconductor, when implementedas such, the pressure in the vessel is preferably maintained at aconstant level. Accordingly, the linked saturation temperature for theprocess medium is generally also known.

Hence, in order to increase the efficiency of the cryogenicrefrigeration system, the outlet of the heat exchanger may be coupled toa recuperation system, a compressor system, a vacuum pump, and/or aliquefaction system, which is configured to provide a constant pressurein the vessel. The matching temperatures of the exhaust gas and thesupply flow at the warm end of the heat exchanger, in particular in thetemperature range between 4.5 and 20 K, facilitates the conversion andrecycling of the evaporated gas in the system. For example, asub-atmospheric evaporated gas may be recuperated and/or a warm, cold,or mixed compression of a sub-atmospheric gas flow may be provided.Accordingly, the outlet of the heat exchanger may be intermittentlycoupled to the supply flow entry of the system, such that a closedcryogenic refrigeration system is provided.

Preferably, the process medium provided upstream of the first pressureregulator is a pressurized fluid, for example helium or nitrogen.However, different process media may be used. The provision of a liquidprocess medium at least has the advantage that the flow parameterswithin the heat exchanger section may be controlled and optimized andimproved heat transfer between the process medium and the heat exchangermay be provided. For example, the supply flow may be configured toprovide the required flow characteristics such as turbulence andboundary layers to increase the heat transfer. The supply pressure ofthe process medium in the conduit is thereby preferably maintained at aconstant value to mitigate pressure fluctuations due to undesirablethermoacoustic oscillations, which e.g. could be caused by safetymechanisms such as safety valves to the 300K-level. Furthermore, byproviding the process medium as a pressurized fluid, the heat capacityof the process medium may be varied by relaxation of the pressurizedliquid in the first pressure regulator and/or adjusting the supply flow.In addition, the first pressure regulator may be configured to reducethe pressure of the process medium to provide a two-phase process mediumflow downstream of the first pressure regulator. For example, thepressure reduction results in a reduced saturation temperature of theprocess medium, such that at least a part of the process medium isconverted from the liquid phase to the gas phase. To adjust the pressureof the process medium, the pressure regulator preferably comprises avalve, expansion valve, and/or turbine. By providing the pressureregulator, both the specific enthalpy of the process medium and the massflow of the liquid phase downstream of the pressure regulator may beadapted to adapt to variable conditions, e.g. due to different heatcapacities at each pressure level and for each physical state.

Preferably, the vessel collects the liquid phase, wherein the vessel isthermally coupled to a load or wherein a load is disposed in thecollected liquid phase of the vessel to provide an isothermal load. Forexample, the vessel may be dimensioned, such that the liquid phaseimmerses a load provided on the bottom of the vessel, e.g. to maximize aheat transfer area between the liquid phase and the load. Alternatively,the load may be thermally coupled to the vessel, for example, by meansof fluid coupling and/or heat conducting surface. By the same token, thevessel may be dimensioned to at least partially enclose a load, whereinthe liquid phase in the vessel may be either collected or circulatedaround at least a portion of the load. Furthermore, although for certainapplications a liquid phase may be preferred, alternatively, the vesselmay be dimensioned to generate and partially retain a sub-atmosphericevaporated gas, which may be used to isothermally cool the load.Preferably, the isothermal load is provided below the saturationtemperature of the process medium at atmospheric pressure. The vesselmay e.g. be configured as a cryostat or cryogenic user such as asuperconductor.

The evaporated gas from the process medium is preferably provided by astate of the two-phase process medium controlled by the pressureregulator, a pressure of the vessel, and the load, wherein the generatedevaporated gas is a sub-atmospheric evaporated gas. Accordingly, thepressure regulator may adiabatically relax the process medium to providea part of the process medium having a gas phase, wherein the state orspecific enthalpy of the process medium downstream of the pressureregulator is dependent on a predefined expansion or pressure relaxationby the pressure regulator and a generally predefined state of the supplyflow upstream of the pressure regulator, which is normally defined bythe regulated constant supply pressure and a temperature slightly abovethe A-temperature, since lower temperatures are generally not reached inthe heat exchanger due to the heat capacity peak around the A-line andthe thermal conductivity increase. The pressure in the vessel isfurthermore preferably remained at a constant level, such that thevessel configuration and pressure cause a further pressure drop of theprocess medium, such that evaporated gas at sub-atmospheric pressure isgenerated. The sudden expansion of the process medium in the vessel mayfurther provide an evaporated gas and flash gas from the liquid phaseresulting from the Joule-Thomson expansion. In addition, the generationof the sub-atmospheric evaporation gas is dependent on the load, whichcauses the liquid phase, preferably provided below the saturationtemperature, to at least partially reach a temperature above thesaturation temperature. The sub-atmospheric evaporated gas maysubsequently enter the inlet of the heat exchanger to cool the supplyflow in the heat exchanger section of the conduit. This at least has theadvantage that the latent heat of the evaporated gas is at the lowestlevel in the system, such that improved heat absorption takes placewithin the heat exchanger. Furthermore, exergetic losses occurringwithin an evaporation heat exchanger are minimized using thesub-atmospheric evaporated gas as a coolant or refrigerant for thesupply flow.

The cryogenic refrigeration system may further comprise a controller andat least one sensor in communication with the controller. Accordingly,the system may comprise at least one temperature sensor arrangedupstream of the pressure regulator and downstream of the heat exchangersection, wherein the controller is configured to control the firstpressure regulator based on the measured value of the at least onetemperature sensor to control the state of the two-phase process medium.Alternatively, or in addition, the system may comprise at least onefilling sensor arranged in the vessel and/or at least one flow sensorarranged downstream of the pressure regulator for measuring a mass flowof a liquid phase of the process medium to the load, wherein thecontroller is configured to control the pressure regulator to controlthe mass flow based on the measured value of the at least one fillingsensor and/or the at least one flow sensor, and/or at least one pressuresensor arranged in communication with the vessel and a compressor systemcoupled to the outlet of the heat exchanger, wherein the controller isconfigured to control the pressure in the vessel by controlling thecompressor system based on the measured value of the at least onepressure sensor. For example, as the temperature and pressure of thesupply flow are generally regulated at a constant level and may hence beconsidered as fixed boundary conditions, a measured temperaturedeviation from a predefined temperature by the temperature sensorarranged downstream of the heat exchanger section and upstream of thepressure regulator may be corrected by accordingly adjusting thepressure regulator to control the state of the process medium, e.g. thespecific enthalpy, downstream of the pressure regulator. As the pressureand the load in the vessel are considered to be constant, a change inthe state of the two-phase process medium hence changes the volume flowof the sub-atmospheric evaporated gas entering the heat exchanger at thecold end. Accordingly, the measured temperature deviation of the processmedium downstream of the heat exchanger section is corrected.

By the same token, a filling sensor may indicate an increased activityof a cryogenic load, such that an increased mass flow of the processmedium to the load is required. Alternatively, or in addition, suchindication may be provided by a flow sensor arranged downstream of thepressure regulator for measuring a mass flow of a liquid phase of theprocess medium to the load. Accordingly, the controller may adjust thepressure regulator to e.g. increase the mass flow according to therequired isothermal load corresponding to the measured value of thefilling sensor and/or the flow sensor. The controller may hencecompensate the discrepancy between the mass flow needed to hold apredefined level in the liquid vessel, e.g. due to an increasedevaporated gas phase provided by the load and a corresponding liquidphase deficit in the vessel, via the pressure regulator.

In addition, a feedback provided by a pressure sensor to the controllermay indicate an undesirable pressure drop or overpressure in the vessel,which is preferred to be maintained at a constant pressure to providecontinuous conditions and a predictable physical impact on the loadcoupled to or provided in the vessel. Accordingly, a compressor systemcoupled downstream of the vessel at the outlet of the heat exchanger maybe adjusted to normalize the pressure of vessel and hence the processmedium and evaporated gas to a tolerable predefined range.

Hence, the controller and the sensor arrangement provide for a feedbackmechanism that provides a means to control the boundary conditions andparameters of the system within a predefined range.

The cryogenic refrigeration system may further comprise a control valvefor controlling the mass flow of the supply flow, which is incommunication with the controller and is arranged in parallel to andupstream of the first pressure regulator, wherein the controller isconfigured to control the mass flow of the supply flow via the controlvalve based on the measured value of the at least one temperaturesensor, filling sensor, and/or flow sensor.

The control valve may hence be adjusted in response to systemfluctuations, e.g. to adjust the liquid phase in the vessel and/or thevolume of evaporated gas provided to the heat exchanger. The controlvalve may e.g. be configured to provide a partial bypassing of thesupply flow to correct for an excess volume flow in the conduit, whereinthe bypass may forward the excess volume flow to adjacent systems or mayrecollect said volume. By the same token, a parallel supply flow maycompensate a deficit of the liquid phase in the vessel and may hence bepartially fed to the supply flow via the parallel control valve.Alternatively, the supply flow may provide a volume flow which slightlyexceeds the required volume flow to compensate for the occurrence of adeficit, wherein the parallel control valve continuously bypasses theexcess supply flow to adjacent systems and does not bypass said excessin case of a detected deficit in the vessel.

For example, while maintaining a constant pressure of the supply flow,the controller may increase the volume and/or flow rate of the supplyflow by accordingly adjusting the control valve, e.g. when the fillsensor indicates a reduced fill status of the liquid phase of theprocess medium in the vessel. Furthermore, the controller may adjust theflow rate of the supply flow, even when the fill status of the liquidphase indicates a normal range during normal operation, but an increasedmass flow of evaporated gas is required. The controller may then controlthe first pressure regulator and the control valve, such that the volumeof sub-atmospheric evaporated gas is increased while the level of theliquid phase of the process medium is remained constant, e.g. byadjusting the currently set value of the first pressure regulator andhence the specific state of the process medium, such that the pressureand hence the enthalpy of the cooled process medium is reduced while atthe same time the volume flow or flow rate of the supply flow isincreased by correspondingly adjusting the control valve. This results,given that the pressure and the load in the vessel are kept constant, inan increased gas phase in the two-phase process medium and a largervolume of sub-atmospheric evaporated gas while the volume of the liquidphase of the process medium collected in the vessel remains essentiallyunchanged.

The controller in the cryogenic refrigeration system may furthermore beconfigured to adjust the first pressure regulator to provide the processmedium downstream of the heat exchanger section of the conduit at atemperature between the lambda point and the saturation temperatureduring normal operation of the cryogenic refrigeration system.Preferably, said temperature range is obtained upstream of the firstpressure regulator, such that a gas phase of the process mediumdownstream of the first pressure regulator comprises a temperaturewithin said range prior to entry into the vessel. The pressure and theload in the vessel are preferably remained constant, while the pressureprovided in the vessel is lower compared with the pressure upstream ofthe pressure regulator. Accordingly, further relaxation of the processmedium in the vessel due to a sudden volume expansion, may result in afurther pressure drop, causing a further reduction of the latent heatand/or temperature of the evaporated gas due to e.g. a Joule-Thomsonexpansion and may hence provide an improved cooling of the supply flowby the heat exchanger. The fixed pressure of the process medium as afixed boundary condition of the supply flow at a temperature between thelambda point and the saturation temperature downstream of the heatexchanger section and upstream of the first pressure regulatorfurthermore ensures that a stable physical state of the process mediumis provided, such that heat transfer fluctuations are minimized.

Furthermore, the system may comprise at least one warm-end temperaturesensor in communication with the conduit and the outlet of the heatexchanger at the warm end of the heat exchanger, wherein the controlleris configured to adjust the evaporative gas flow based on a temperaturedifference measured by the at least one warm-end temperature sensor bycontrolling the pressure regulator.

While the temperature of the supply flow at the warm end of the heatexchanger is generally considered a fixed boundary condition, thetemperature measured by the sensor at the outlet at the warm of the heatexchanger may be dependent e.g. on the heat exchanger efficiency orprovided cooling of the supply flow and hence the state of the processmedium upstream of the pressure regulator as well as the cryogenic loador mass flow. Accordingly, to minimize the temperature differencedetected at the warm end of the heat exchanger, the controller mayincrease the sub-atmospheric evaporated gas flow and/or the mass flowtowards the load, as outlined in the above, e.g. by adjusting thepressure regulator and/or the control valve, preferably based on ameasured temperature of the process medium by a temperature sensorprovided upstream of the pressure regulator and downstream of the heatexchanger section.

In addition, the achieved temperature range of the liquid phase may notonly be used to provide an isothermal load, but may also provide aliquid phase to be implemented in systems configured for studying e.g.molecular interactions and fluid characteristics, for example, to studythe transition from helium-1 to helium 2 at the lambda point and thesuperfluidity or viscosity behavior of helium at supercriticaltemperatures.

The heat exchanger of the cryogenic refrigeration system may beconfigured as a plurality of heat exchanging modules, which are arrangedin parallel and/or in series to the conduit. Preferably, a secondpressure regulator in fluid communication with the conduit is arrangedbetween each serially arranged heat exchanging module.

For example, the heat exchanger may comprise two heat exchanging modulesthat are arranged in series to the conduit, wherein between said heatexchanger modules, a second pressure regulator, e.g. a valve orexpansion turbine, is arranged and in fluid communication with theconduit. This at least has the advantage that the supply flow aftercooling by the first heat exchanger module may be throttled by anadditional pressure regulator to an intermediate pressure level prior tocooling by the second heat exchanger module, thereby increasing the heatcapacity and providing a gradual relaxation of the process medium. Atthe same time, the temperature level on the warm end of the first heatexchanger module may be increased with regard to a single heat exchangerconfiguration. Accordingly, the provision of a plurality of heatexchanger modules may further increase the efficiency of the process.

According to a further aspect of the invention, a method for providing acryogenic refrigeration in a cryogenic refrigeration system issuggested, wherein the method comprises the steps of

-   -   providing a supply flow of a process medium in a conduit;    -   cooling the supply flow in a counter flow heat exchanger;    -   reducing the pressure of the supply flow by means of a pressure        regulator; and    -   receiving the supply flow in a vessel, wherein an evaporated gas        flow from the process medium is used by the heat exchanger to        cool the supply flow,        wherein the cooling of the supply flow is provided free of any        evaporating liquid phase.

Accordingly, the cooling of the supply flow or the process medium occursby the gas flow with low enthalpy that has evaporated prior to entryinto the heat exchanger. Hence, no liquid phase enters the heatexchanger, such that, contrary to an evaporating heat exchanger, noliquid phase is evaporated within the heat exchanger. This isparticularly advantageous when using liquid helium, such that the systemdoes not require an evaporating heat exchanger and phase separator atthe 4.5 K level and furthermore no recirculation of flash gas orevaporated helium on the atmospheric pressure occurs. In addition,smaller equipment such as compressors and heat exchangers may beprovided, such that the dimensions of the system may be reduced.

Furthermore, the method may comprise that a temperature factor of theevaporated gas at a warm end of the heat exchanger relative to theprocess medium of the supply flow at the warm end of the heat exchangeris provided by the heat exchanger, which is larger than 0.9 duringnormal operation of the cryogenic refrigeration system. Preferably, saidtemperature factor is larger than 0.98, such that temperaturedifferences between the evaporated gas at the warm end of the heatexchanger relative to the process medium of the supply flow at the warmend of the heat exchanger are minimal and/or negligible, thereby notaffecting the system.

Such temperature factor is possible since the system does not require anupstream evaporating heat exchanger, which generally providestemperatures of the process medium after passing the evaporating heatexchanger fixed at around e.g. 4.6K for helium, wherein the mass flowsin a steady state process at the cold and warm ends are generally equaland constant. In contrast, the cold counter flow heat exchanger mayprovide the supply flow and the evaporating gas at a higher temperaturelevel at the warm end of the heat exchanger having an increased heatcapacity, such that temperature differences may be minimized.

Alternatively, or in addition, a temperature of the evaporated gas ismatched to a temperature of the process medium at a warm end of the heatexchanger during normal operation of the cryogenic refrigeration systemprovided by an NTU configuration of the heat exchanger.

The implementation of a heat exchanger comprising the required NTU atleast has the advantage that the system may be thermodynamicallyoptimized while certain variables, e.g. heat exchanger parameters andboundary conditions, are not required or need not be known. Accordingly,the NTU configuration provides an alternative to the LMTD configurationto provide a thermally efficient cryogenic refrigeration system.

As outlined in the above, matching temperatures or a minimal temperaturedifference by implementation of a heat exchanger comprising the requiredNTU at least has the advantage that exergetic losses are reduced duringnormal operation. As such, the occurrence of irreversibilities in thesystem and the power input to the process are likewise reduced.

The method furthermore preferably provides that the supply flowcomprises a pressurized liquid, preferably liquid helium, whereinreducing the pressure of the supply flow by the pressure regulatorprovides a two-phase process medium flow downstream of the pressureregulator and wherein the evaporated gas in the vessel is provided atsub-atmospheric pressure. Providing the process medium as a pressurizedliquid may facilitate the heat transfer in the heat exchanger section ofthe conduit and the handling of the process medium, e.g. providing thesupply flow.

Preferably, the cooling of the supply flow provides the process mediumbetween the lambda point and the saturation temperature downstream ofthe heat exchanger section of the conduit. As outlined in the above,such temperature range and having a fixed pressure as a boundarycondition ensures that a stable physical state of the process medium ismaintained and may hence reduce the occurrence of fluctuations in thesystem. At the same time, releasing the pressure of the process mediumdownstream of the heat exchanger section may then result in differentphysical states of the process medium, such that e.g. both a liquidphase and gas phase are obtained.

The method may furthermore provide a cryogenic refrigeration of a load.Accordingly, the vessel may collect the liquid phase of the processmedium to refrigerate a thermally coupled load or a load disposed in theliquid phase of the process medium in the vessel, to provide anisothermal load.

To further optimize the efficiency of the cryogenic refrigerationmethod, the cooling of the supply flow may occur in series or inparallel by means of a plurality of heat exchanger modules arranged inseries or in parallel. In such configuration, the pressure of the supplyflow is preferably reduced between each serially arranged heat exchangermodule by means of a second pressure regulator. The throttling of theprocess medium between the heat exchanger modules has the advantage thatan intermediate pressure level is obtained and the heat capacityincreased while furthermore a gradual relaxation of the process mediumis provided. In addition, the serial cooling of the process mediumprovides that the temperature level on the warm end of the heatexchanger arrangement may be increased, such that the efficiency of theprocess is increased.

BRIEF DESCRIPTION OF THE DRAWINGS

The present disclosure will be more readily appreciated by reference tothe following detailed description when being considered in connectionwith the accompanying drawings in which:

FIG. 1 is a schematic view of a heat exchanger, a vessel, and a pressureregulator in a cryogenic refrigeration system;

FIG. 2 is a schematic view of the embodiment according to FIG. 1configured to provide the process medium in predefined physical states;

FIG. 3A is a schematic cross-sectional view of a tubular heat exchanger;

FIG. 3B is a schematic top view of the tubular heat exchanger accordingto FIG. 3A seen from the cold end of the heat exchanger;

FIG. 4 is a schematic view of a cryogenic refrigeration system having acontroller and a load;

FIG. 5 is a schematic view of the cryogenic refrigeration systemaccording to FIG. 4 with a further controller configuration;

FIG. 6A is a schematic view of a cryogenic refrigeration system having aserial heat exchanger and pressure regulator arrangement;

FIG. 6B is a schematic view of the cryogenic refrigeration systemaccording to FIG. 6A, comprising a further parallel heat exchangerarrangement.

DETAILED DESCRIPTION OF PREFERRED EMBODIMENTS

In the following, the invention will be explained in more detail withreference to the accompanying figures. In the Figures, like elements aredenoted by identical reference numerals and repeated to descriptionthereof may be omitted in order to avoid redundancies.

In FIG. 1 a cryogenic refrigeration system 1 is schematically shown inoperation using a process medium. In order to provide refrigeration, asupply flow 10 of a process medium is provided in the conduit 2.Although the process media may comprise various compounds and mayfurthermore be provided in different physical states, the process mediumin the exemplary embodiment according to FIG. 1 comprises pressurizedliquid helium. The liquid helium is hence at the pressure aboveatmospheric pressure, preferably between 1.5 and 10 bar, more preferablybetween 1.5 and 8.0 bar.

All features of the system 1 and in particular the conduit 2 arethermally isolated, such that the amount of heat entering and leavingthe system 1 is considered to be zero or negligible. The cryogenicrefrigeration system 1 comprises a counter flow heat exchanger 3, whichis thermally coupled to a heat exchanger section 2A of the conduit 2,such that the supply flow 10 is cooled by means of the counter flow heatexchanger 3. After cooling by the heat exchanger 3 the supply flow 10arrives at a first pressure regulator 4, which is in fluid communicationwith the conduit 2 and is arranged downstream of heat exchanger section2A of the conduit 2. In this context the term “downstream” refers to thesupply flow 10 provided in the conduit 2 and in relation to the initialentry of the supply flow 10 into the system 1. Accordingly, the entry ofthe supply flow 10 into the system 1 occurs upstream of the heatexchanger section 2A.

The first pressure regulator 4 is provided as an expansion valve orvalve arrangement. By means of the first pressure regulator 4 thepressure of the process medium in the supply flow 10 is reduced to apressure slightly above atmospheric pressure, e.g. 1.05 to 1.2 bar. Thesupply flow 10 then flows into a vessel 5, which is in fluidcommunication with the conduit 2 and is hence arranged downstream of thefirst pressure regulator 4. Although the fluid communication between thefirst pressure regulator 4 and the vessel 5 is depicted in FIG. 12 tocomprise a conduit, e.g. an outlet of the first pressure regulator 4and/or a corresponding inlet of the vessel 5, the fluid communicationmay also be provided by coupling a downstream end of the first pressureregulator 4 directly to a corresponding opening or coupling element ofthe vessel 5.

The vessel 5 comprises a constant pressure, which is lower than thepressure upstream of the vessel 5, and is configured to collect a liquidphase and provide an evaporated gas from the process medium. Theevaporated gas is generated depending on the state of the process mediumdownstream of the first pressure regulator 4, e.g. the specificenthalpy, any boundary activities or implementations of the vessel 5,e.g. a load (not shown), and the pressure in the vessel, which isremained constant. Due to a sudden volume increase in the vessel 5compared to the volume of the process medium downstream of the firstpressure regulator 4, the process medium is further relaxed downstreamof the first pressure regulator 4. For example, the vessel 5 is sizedand dimensioned to promptly expand the process medium. The sudden volumeincrease of the process medium in the vessel 5 results in a rapidpressure reduction of the process medium, such that a gas phase or flashgas is generated, which comprises a sub-atmospheric pressure, i.e. below1.0 bar. In this Joule-Thomson expansion, the temperature of thesub-atmospheric evaporated gas may remain constant or is slightlyreduced while the latent heat of the evaporated gas is reduced. Inaddition, as outlined in the above, an implementation of the vessel 5may cause the liquid phase in the vessel also to provide an evaporatedgas. Accordingly, the sub-atmospheric evaporated gas 12 is then providedto an inlet 34 of the heat exchanger 3 to serve as a coolant orrefrigerant for the supply flow 10 of the process medium. The inlet 34of the heat exchanger 3 may be either directly coupled to the vessel 5or may be fluidly connected to an outlet of the vessel 5 by means of aconduit or tube section.

As the latent heat and temperature of the sub-atmospheric evaporated gas12 are considered to be at its lowest in the system 1 at the inlet 34 ofthe heat exchanger 3, this region is considered the cold end 30 of theheat exchanger 3. During cooling of the supply flow 10 of the processmedium in the heat exchanger section 2A by the sub-atmosphericevaporated gas 12 in the heat exchanger 3, the sub-atmosphericevaporated gas 12 absorbs heat from the supply flow 10 of the processmedium, such that the outlet 36 of the heat exchanger 3 is considered tobe a warm end 32 of the heat exchanger 3. Accordingly, thesub-atmospheric evaporated gas 12 flows from an inlet 34 at the cold end30 of the heat exchanger 3 to an outlet 36 at a warm end 32 of the heatexchanger 3, thereby absorbing heat from the supply flow 10 of theprocess medium and transitioning from a cold sub-atmospheric evaporatedgas 12 to a warm sub-atmospheric evaporated gas 12, and leaves thesystem 1 at the outlet 36 as an exhaust gas 14.

Although the cryogenic refrigeration system 1 requires a normalizationand stabilization of the temperatures in the system 1 during start up oran initial phase of operation, the temperature of the process medium atvarious points or locations in the system 1 is considered to be constantand predictable during normal operation. Accordingly, the process mediumin the vessel 5 may be used to provide isothermal conditions, e.g. anisothermal load (not shown).

The conduit 2 is free of any evaporation heat exchanger upstream of theheat exchanger section 2A of the conduit 2. Accordingly, by providing acold counter flow heat exchanger 3 comprising an evaporated gas flowwith low specific enthalpy, the system does not require an evaporator toprecool the supply flow 10. Furthermore, the cold counter flow heatexchanger 3 may provide the supply flow 10 and the evaporated gas 12 ata higher temperature level at the warm end 32 of the heat exchanger 3having an increased heat capacity, such that temperature differences maybe minimized.

In particular, the heat exchanger 3 of the system 1 is configured suchthat during normal operation the temperature of the exhaust gas 14matches the temperature of the supply flow 10 of the process medium atthe warm end 32 of the heat exchanger 3. The term “matches” here is tobe understood to also include minimal differences, e.g. up to 0.5 K,preferably between 0.05 and 0.2 K. This matching minimal difference ofsaid temperatures is achieved by the configuration of the heat exchanger3, wherein the corresponding NTU or the heat transfer rate isaccordingly adapted. For example, the area of the heat exchanger 3, e.g.the heat transfer area or length of the heat exchanger 3 may be sizedand dimensioned to provide the corresponding temperature range, whereinat least the mass flow and the heat capacity values at varioustemperatures of the process medium are considered to be known. Forexample, the heat transfer area of the heat exchanger 3 may be sized toprovide the required NTU to provide a sufficient cooling of the processmedium such that the process medium downstream of the heat exchangersection 2A of the conduit 2 and upstream of the first pressure regulator4 is provided above the lambda point at a temperature of 2.14 to 2.40 Kwhile at the same time providing a temperature of the exhaust gas 14matching the temperature of the process medium upstream of the heatexchanger section 2A at the warm end 32 of the heat exchanger 3 between4.5 and 20 K or even higher, preferably around 12 K. The correspondingNTU of the heat exchanger 3 may hence be optimal for said temperatureranges of liquid helium. However, the NTU may be adapted for othertemperature ranges and/or compounds and may furthermore provide anexcess to accommodate system fluctuations or variating needs, e.g. of aload to be cooled by the system 1.

The cryogenic refrigeration system 1 according to FIG. 2 largelycorresponds to the embodiment depicted in FIG. 1. Again, the processmedium is provided by a supply flow 10 in the conduit 2 and is cooled bythe heat exchanger 3 as described in the above. In addition, the heattransfer area of the heat exchanger 3 is adapted to provide a heattransfer rate, which provides a cooling of the process medium resultingin a cooled process medium 11, e.g. comprising a temperature just abovethe lambda point and below the saturation temperature of thecorresponding pressure of the supply flow 10, for example between 2.14and 2.40 K. The pressure of the cooled process medium 11 is then reducedby the first pressure regulator 4 or expansion valve, to obtain atwo-phase process medium 13. In other words, the pressurized liquidhelium in the supply flow 10 is first cooled by the heat exchanger 3 toa predetermined temperature and is subsequently depressurized to providea process medium comprising a liquid and gas phase.

The configuration of the vessel 5 is such that the liquid phase 15 ofthe two-phase process medium 13 is collected upon entry into the vessel5 while at the same time the configuration, e.g. the dimensioning andthe constant pressure in the vessel 5, causes the generation ofsub-atmospheric evaporated gas 12, depending on the respective state ofthe two-phase process medium 13. The sub-atmospheric evaporated gas 12then flows into the heat exchanger 3 via an inlet 34 at a cold end ofthe heat exchanger 3 to cool the supply flow 10. The sub-atmosphericevaporated gas 12 leaves the heat exchanger 3 at a warm end 32 of theheat exchanger 3 and exits the system 1 via an outlet 36 as exhaust gas14.

Accordingly, the cryogenic refrigeration system 1 according to FIG. 2 isoptimized to both provide a sufficient cooling of the supply flow 10 bythe sub-atmospheric evaporated gas 12 and a sufficient amount of theliquid phase 15 of the process medium at a required temperature, e.g.for further refrigeration requirements, by means of a correspondingdepressurization of the supply flow 10 to provide a two-phase processmedium 13, a configuration and constant pressure of the vessel 5, and aconfiguration of the heat exchanger 3, e.g. by a corresponding NTU orheat transfer rate.

In FIGS. 3A and 3B the counter flow heat exchanger 3 is schematicallyshown in further detail. The process medium is provided by means of thesupply flow 10 in the conduit 2. The heat exchanger 3 comprises a tubeshape, which surrounds the circumferential area of the conduit 2 forminga heat exchanger section 2A. Although the heat exchanger 3 is depictedcomprising a cylindrical form and fully surrounding the conduit 2, othershapes and configurations are possible. However, in any case the NTU ofthe heat exchanger 3 is predefined to accordingly cool the supply flow10 and minimize a temperature difference of the exhaust gas 14 and thesupply flow 10 at the warm end 32 of the heat exchanger 3.

As shown in FIG. 3A, the heat exchanger section 2A of the conduit 2linearly traverses the heat exchanger 3 from the warm end 32 to the coldend 30 of the heat exchanger 3 and comprises a substantially straightconfiguration. However, other configurations that increase the heattransfer rate or are thermodynamically efficient are possible, forexample, a meandering, sinusoidal, or coiled shape of the conduit 2.While traversing the heat exchanger 3 the supply flow 10 is cooled bythe heat exchanger 3 by means of the sub-atmospheric evaporated gas 12entering the heat exchanger 3 at the cold end 30 via an inlet 34.

The cooling of the supply flow 10 is provided by the sub-atmosphericevaporated gas 12, which is distributed through the heat exchanger 3 bymeans of a spirally formed heat exchanger element 38. Accordingly, thespirally formed heat exchanger element 38 traverses the heat exchanger 3in a counter flow direction of the conduit 2, wherein thesub-atmospheric evaporated gas 12 absorbs the heat from the supply flow10 provided in the thermally coupled heat exchanger section 2A of theconduit 2, either through direct contact or thermal coupling by means ofa heat conducting material. At the warm end 32 of the heat exchanger 3the sub-atmospheric evaporated gas 12 then exits the heat exchanger 3via an outlet 36 as an exhaust gas 14.

The inlet 34 and the outlet 36 of the heat exchanger 3 are arranged inparallel and adjacent to the conduit 2 at the cold end 30 and the warmend 32 of the heat exchanger 3, respectively. This configuration is alsoshown in FIG. 3B, which shows the heat exchanger 3 from a perspective inflowing direction of the sub-atmospheric evaporated gas 12 and incounter flowing direction of the cooled process medium 11 at the coldend 30 of the heat exchanger 3. Although the conduit 2 and the inlet 34of the heat exchanger 3 are arranged adjacently in a verticalorientation, any orientation that is perpendicular to an extendingdirection of the heat exchanger 3 or the spirally formed heat exchangerelement 38 or a substantial lateral arrangement may be provided. By thesame token, the spirally formed heat exchanger element 38 may bearranged adjacent to the conduit 2 within the heat exchanger 3 toprovide a direct heat transfer between the spirally formed heatexchanger element 38 and the conduit 2. Accordingly, the heat exchanger3 may be alternatively dimensioned to comprise a smaller size in aradial direction.

However, other configurations of the heat exchanger 3 may be provided.For example, the heat exchanger 3 may be configured as a plate fin heatexchanger, e.g. for larger systems or plants, or as a coil finned tubeheat exchanger, e.g. for smaller systems or plants. In a plate fin heatexchanger, the heat exchanger comprises a plurality of compartments thatare arranged adjacently and in countercurrent orientation to each otherand wherein said compartments comprise either the sub-atmosphericevaporated gas or the supply flow. When implementing the heat exchanger3 as a coil finned tube heat exchanger on the other hand, thesub-atmospheric evaporated gas may be guided along the conduit 2comprising the supply flow 10 in a coiled fashion, wherein the coiledarrangement furthermore comprises a plurality of loop sections thatextend radially outward, thereby defining a plurality of fins.

A further embodiment of the cryogenic refrigeration system 1 is shown inFIG. 4. FIG. 4 essentially corresponds to the system 1 according to FIG.2, such that like features and functions are not discussed in furtherdetail. The system 1 comprises a controller 7, which is in communicationwith the first pressure regulator 4 and is configured to control thefirst pressure regulator 4 in order to relax or expand the cooledprocess medium 11 to provide a two-phase process medium 13 downstream ofthe first pressure regulator 4. To appropriately adjust the pressure ofthe cooled process medium 11, the controller 7 is in communication witha temperature sensor 70 which is in communication with the conduit 2 andthe outlet 36 of the heat exchanger 3 at the warm end 32 of the heatexchanger 3. Said sensor 70 hence provides an actual temperature of thesupply flow 10 entering the system 1 and the exhaust gas 14 exiting thesystem 1 via the outlet 36. The measured values of the sensor 70 areprovided to the controller 7, wherein the controller 7 controls thefirst pressure regulator 4 based at least on the measured values of thesensor 70, the state of the two-phase process medium 13, and thepressure in the vessel 5.

Although the system 1 is generally designed for specific boundaryconditions and the status of the system 1 is maintained constant, theprovision of the controller 7 and the temperature sensor 70 allow thesystem 1 to react to or prevent minor fluctuations in the system 1, e.g.by adjusting the volume flow of the sub-atmospheric evaporated gas 12.The volume flow of sub-atmospheric evaporated gas 12 is dependent on thestate of the two-phase process medium 13 and the pressure in the vessel5, which is maintained at a constant level by a compressor (not shown)in communication with the vessel 5 at a downstream end, e.g. downstreamof the outlet 36. As both the temperature and the pressure of the supplyflow 10 are fixed boundary conditions and the cooling efficiency of theheat exchanger 3, and therefore the state of the cooled process medium11, is generally known, the state or specific enthalpy of the two-phaseprocess medium may be controlled by adjusting the pressure regulator 4.For example, the controller 7 may adjust the first pressure regulator 4to further reduce the pressure of the cooled process medium 11, when anundesirable temperature difference between the exhaust gas 14 and thesupply flow 10 is measured, e.g. when the measured temperature of theexhaust gas 14 is higher than the temperature of the supply flow 10,such that the two-phase process medium 13 is relaxed and/or the gasphase is increased and hence, at a constant vessel pressure, a largervolume flow of sub-atmospheric evaporated gas 12 is provided to the heatexchanger 3. Accordingly, an improved cooling of the supply flow 10 maybe provided while at the same time absorbed heat in the sub-atmosphericevaporated gas 12 levels out the temperature difference between theexhaust gas 14 and the supply flow 10 at the warm end 32 of the heatexchanger 3.

Provided in the liquid phase 15 of the process medium collected in thevessel 5 is a load 6. The load also affects the volume flow ofsub-atmospheric evaporated gas 12 as, depending on the activity of theload 6, the liquid phase 15 may partially attain a temperature above thesaturation temperature and hence enter the gas phase. In order tomaintain an isothermal load 6, the controller 7 may hence accordinglyadjust the first pressure regulator 4 to e.g. compensate for a loss ofliquid phase 15. For example, the controller 7 may adjust the pressureand hence the specific enthalpy of the two-phase process medium 13 bycontrolling the first pressure regulator 4 to increase the liquid phase15 of the two-phase process medium 13 to be collected in the vessel 5and to compensate for an increased amount of sub-atmospheric evaporatedgas 12 and a loss of the liquid phase 15 in the vessel 5. By the sametoken, the change in mass flow to the load 6 may be detected by a changein temperature, which is measured by the temperature sensor 70 and maybe provided to the controller 7 as feedback.

In addition to the temperature sensor 70, the embodiment according toFIG. 5 comprises a fill sensor 72 and a pressure sensor 74 disposed inthe vessel 5 that are in communication with the controller 7.Accordingly, the controller 7 controls the first pressure regulator 4 byadjusting the pressure of the cooled process medium 11 based on a fillstatus measured by the fill sensor 72 in the vessel 5. For example, anincrease in the activity of the load 6 may reduce the fluid level of theliquid phase 15 of the process medium, which is detected by the fillsensor 72 and indicates to the controller that a deficit of the liquidphase 15 is present in the system 1. The controller 7 may then controlthe first pressure regulator 4 to accordingly adjust the state of thetwo-phase process medium 13 and hence the liquid phase 15 provided tothe vessel 5.

In addition, the controller 7 is in communication with a control valve20, which is arranged in parallel to and upstream of the pressureregulator 4. The control valve 20 is configured as a three-way-valve andconnects the conduit 2 to a parallel system. Should the fill sensor 72indicate a deficit or excess of the liquid phase 15 of the processmedium in the vessel 5, the controller 7 may control the control valve20 to accordingly adjust the mass flow while retaining a constantpressure and temperature of the supply flow. Alternatively, or inaddition, such indication may be provided by a flow sensor 76 incommunication with the controller 7 and provided downstream of thepressure regulator 4 and indicating a mass flow to a load 6

The pressure in the vessel 5 is furthermore maintained at a constantlevel by a compressor (not shown) in communication with the vessel 5 ata downstream end, e.g. downstream of the outlet 36. The pressure in thevessel 5 is measured by the pressure sensor 74. Should a pressuredeviation from a predefined range or threshold occur, said pressuresensor 74 provides a feedback to the controller 7, which accordinglyadjusts the pressure via the downstream compressor.

Furthermore, a temperature sensor 70 is provided, which is arrangeddownstream of the heat exchanger section 2A and upstream of the pressureregulator 4 and is in communication with the controller 7. As thetemperature and pressure of the supply flow 10 are generally regulatedat a constant level and may hence be considered as fixed boundaryconditions, a measured temperature deviation from a predefinedtemperature may be corrected by accordingly adjusting the pressureregulator 4 to control the state of the process medium, e.g. thespecific enthalpy, downstream of the pressure regulator 4. As thepressure and the load 6 in the vessel 5 are considered to be constant, achange in the state of the two-phase process medium 13 hence changes thevolume flow of the sub-atmospheric evaporated gas 12 entering the heatexchanger 3 at the cold end 30. Accordingly, the measured temperaturedeviation of the process medium downstream of the heat exchanger section2A is reduced.

Although the load 6 may be disposed in the liquid phase 15 of theprocess medium in the vessel 5, the load 6 may also be provided outsideof the vessel 5, as depicted in FIG. 5. The volume flow entering andexiting the vessel 5 is hence not affected by the dimensions of the load6 while a thermal coupling between the vessel 5 and the load 6 providesa similar refrigeration of the load 6, e.g. to provide an isothermalload 6. The thermal coupling may be provided by either a direct contactbetween the outer surface of the vessel 5 and the load 6 or by means ofe.g. a fluid coupling such as a check valve.

The heat exchanger 3 may comprise various configurations to provide therequired temperature factor at the warm end of the heat exchanger, e.g.by a corresponding NTU or heat transfer rate. For example, the heatexchanger 3 may comprise a plurality of counter flow heat exchangermodules 3A, 3B, 3C, which are arranged in series and/or in parallel, asshown in the embodiments according to FIGS. 6A and 6B. In FIG. 6A theheat exchanger comprises two heat exchanging modules 3A and 3C that arearranged in series. The serial heat exchanger modules 3A, 3C are fluidlycoupled to each other and thermally coupled with the conduit 2,comprising the process medium.

In operation, the sub-atmospheric evaporated gas 12 enters the secondserial heat exchanger module 3C at a cold end 30 and traverses said heatexchanger module 3C, thereby absorbing heat from the process medium inthe conduit 2. The sub-atmospheric evaporated gas exiting the secondserial heat exchanger module 3C hence comprises a different latent heatand/or temperature compared with the sub-atmospheric evaporated gas 12provided in the inlet 34 and is hence considered a warmedsub-atmospheric evaporated gas 17. The warmed sub-atmospheric evaporatedgas 17 then enters the first serial heat exchanger module 3A and exitsthe system 1 as an exhaust gas 14 at the warm end 32 via an outlet 36.While the warmed sub-atmospheric evaporated gas 17 absorbs heat in thefirst serial heat exchanger module 3A, the process medium in the supplyflow 10 is accordingly cooled, such that the process medium in theconduit 2 arriving at the second serial heat exchanger module 3C isconsidered to be a subcooled process medium 16. Subsequent cooling ofthe subcooled process medium 16 by the second serial heat exchangermodule 3C then results in the cooled process medium 11 downstream of thesecond serial heat exchanger 3C.

The system 1 furthermore comprises a pressure regulating arrangementcomprising a first pressure regulator 4A and a second pressure regulator4B that are in fluid communication with the conduit 2. The firstpressure regulator 4A is arranged downstream of the second serial heatexchanger module 3C and upstream of the first pressure regulator 4A toadjust a pressure of the process medium 11 provide a two-phase processmedium 13 downstream of the first pressure regulator 4A. The secondpressure regulator 4B is arranged between the first and second heatexchanger modules 3A, 3C. this arrangement provides that the pressure ofthe process medium or the pressurized liquid may be adjusted or reducedafter subcooling of the process medium and prior to the cooling by thesecond heat exchanger module 3C to provide the cooled process medium 11,wherein the subcooled process medium 16 may be provided as a liquid oras a two-phase process medium. Accordingly, the system 1 is configuredto optimally use the different heat capacity values of the processmedium for different temperatures and pressures, thereby providing anNTU of the heat exchanger to match the temperatures of the exhaust gas14 and the supply flow 10 at the warm end 32.

A combination of a parallel and serial arrangement of counter flow heatexchanger modules is shown in FIG. 6B. In addition to the first andsecond heat exchanger modules 3A, 3C, the system 1 comprises a parallelheat exchanger module 3B, such that the first serial heat exchangermodule 3A and the parallel heat exchanger module 3B are arranged inparallel. In order to provide such arrangement of the cryogenicrefrigeration system 1, the vessel 5 is fluidly coupled via an inlet 34to a cold end 30 of the second heat exchanger module 3C to provide thesub-atmospheric evaporated gas 12 exiting the vessel 5. After traversingthe second heat exchanger module 3C, the warmed sub-atmosphericevaporated gas is then divided or split into a first and second parallelwarmed sub-atmospheric evaporated gas 17A, 17B and introduced into thefirst serial exchanger module 3A and the parallel heat exchanger module3B, respectively, using parallel fluid couplings. The warmedsub-atmospheric evaporated gas 17A, 17B subsequently exits therespective first serial exchanger module 3A and the parallel heatexchanger module 3B as a first and second exhaust gas 14A, 14B,respectively, wherein the first and second exhaust gas 14A, 14B arecoupled to the outlet 36 at the warm end 32 and are combined to provideexhaust gas 14 exiting the system 1 via the outlet 36.

In order to provide the parallel cooling, the conduit 2 is split intotwo parallel sections that are thermally coupled to the parallelexchanger modules 3A, 3B at a point just before the first serialexchanger module 3A and the parallel heat exchanger module 3B. Theparallel heat exchanger modules 3A, 3B hence provide a subcooling of theprocess medium as described in further detail for the embodimentaccording to FIG. 6A. The parallel sections of the conduit 2 are thenmerged again downstream of the parallel heat exchanger modules 3A, 3Band prior to entry into the second pressure regulator 4B. Downstream ofthe second pressure regulator 4B the process medium is further cooled bythe second serial heat exchanger module 3C and passes the first pressureregulator 4A prior to entry into the vessel 5, as described in relationto FIG. 6A.

According to the embodiment of FIG. 6B, the second serial heat exchangermodule 3C comprises a tube shape surrounding the circumference of theconduit 2, while the parallel heat exchanger modules 3A, 3B are depictedto be thermally coupled to the parallel sections of the conduit 2 in anadjacent manner. However, configurations other than those depicted arepossible, e.g. a plurality of tubular heat exchanger modules and/or heatexchanger modules only partially surrounding the circumference of theconduit 2 may be provided. Furthermore, the conduit sections and thefluid couplings are adjacently arranged to each other to both increasethermal efficiency and reduce the dimensions and size of the system 1.However, it will be understood that other configurations, wherein e.g.the conduit sections and the fluid couplings are at least partiallyspace apart, may also be provided. In particular, further possibleconfigurations of the heat exchanger as described in view of FIGS. 3Aand 3B, i.e. plate fin heat exchanger modules or coil finned tube heatexchanger modules, may also be implemented.

It will be obvious for a person skilled in the art that theseembodiments and items only depict examples of a plurality ofpossibilities. Hence, the embodiments shown here should not beunderstood to form a limitation of these features and configurations.Any possible combination and configuration of the described features canbe chosen according to the scope of the invention.

LIST OF REFERENCE NUMERALS

-   1 Cryogenic refrigeration system-   10 Supply flow of a process medium-   11 Cooled process medium-   12 Sub-atmospheric evaporated gas-   13 Two-phase process medium-   14 Exhaust gas-   14A First parallel exhaust gas-   14B Second parallel exhaust gas-   15 Liquid phase of process medium-   16 Subcooled process medium-   17 Warmed sub-atmospheric evaporated gas-   17A First parallel warmed sub-atmospheric evaporated gas-   17B Second parallel warmed sub-atmospheric evaporated gas-   2 Conduit-   2A Heat exchanger section-   20 Control valve-   3 Counter flow heat exchanger-   3A First serial counter flow heat exchanger module-   3B Parallel counter flow heat exchanger module-   3C Second serial counter flow heat exchanger module-   30 Cold end of heat exchanger-   32 Warm end of heat exchanger-   34 Inlet-   36 Outlet-   38 Spirally formed heat exchanger element-   4 First pressure regulator-   4A First pressure regulator-   4B Second pressure regulator-   5 Vessel-   6 Load-   7 Controller-   70 Temperature sensor-   72 Fill sensor-   74 Pressure sensor-   76 Flow sensor

1. Cryogenic refrigeration system (1), comprising: a conduit (2)configured to provide a supply flow (10) of a process medium; a counterflow heat exchanger (3), which is thermally coupled to a heat exchangersection (2A) of the conduit (2) and comprises an inlet (34) at a coldend (30) of the heat exchanger (3) and an outlet (36) at the warm end(32) of the heat exchanger (3); a first pressure regulator (4), which isin fluid communication with the conduit (2) and is arranged downstreamof the heat exchanger section (2A); and a vessel (5), which is in fluidcommunication with the conduit (2) and is arranged downstream of thefirst pressure regulator (4), wherein the vessel (5) is in fluidcommunication with the inlet (34) of the heat exchanger (3) and isconfigured to provide an evaporated gas flow from the process medium tothe inlet (34) of the heat exchanger (3), wherein the conduit (2) isfree of any evaporation heat exchanger upstream of the heat exchangersection (2A) of the conduit (2).
 2. Cryogenic refrigeration system (1)according to claim 1, wherein the heat exchanger (3) is configured toprovide a temperature factor of the evaporated gas at the warm end (32)of the heat exchanger (3) relative to the process medium of the supplyflow (10) at the warm end (32) of the heat exchanger (3) larger than0.9, preferably larger than 0.98, during normal operation of thecryogenic refrigeration system (1); and/or the heat exchanger (3)comprises an NTU configured to match a temperature of the evaporated gaswith a temperature of the process medium at the warm end (32) of theheat exchanger (3) during normal operation of the cryogenicrefrigeration system (1).
 3. Cryogenic refrigeration system (1)according to claim 2, wherein the temperature factor and/or the NTU isprovided by a heat transfer area of the heat exchanger (3), preferablyby a length of the heat exchanger, wherein the heat exchanger (3) ispreferably of a finned tube shape, coiled shape, and/or fin shape and atleast partially surrounds a circumference of the conduit (2). 4.Cryogenic refrigeration system (1) according to claim 1, wherein thatthe outlet (36) of the heat exchanger (3) is coupled to a recuperationsystem, a compressor system, a vacuum pump, and/or a liquefactionsystem, which is configured to provide a constant pressure in the vessel(5).
 5. Cryogenic refrigeration system (1) according to claim 1, whereinthat the process medium provided upstream of the first pressureregulator (4) is a pressurized liquid, preferably liquid helium orliquid nitrogen, wherein the first pressure regulator (4) is configuredto reduce the pressure of the process medium to provide a two-phaseprocess medium (13) flow downstream of the first pressure regulator (4),wherein the first pressure regulator (4) preferably comprises a valve,expansion valve, and/or turbine.
 6. Cryogenic refrigeration system (1)according to claim 5, wherein the vessel (5) collects the liquid phase(15), wherein the vessel (5) is thermally coupled to a load (6) orwherein a load (6) is disposed in the collected liquid phase (15) of thevessel (5) to provide an isothermal load (6).
 7. Cryogenic refrigerationsystem (1) according to claim 6, wherein the evaporated gas from theprocess medium is provided by a state of the two-phase process medium(13) controlled by the pressure regulator (4), a pressure of the vessel(5), and the load (6), wherein the evaporated gas is a sub-atmosphericevaporated gas (12).
 8. Cryogenic refrigeration system (1) according toclaim 6, wherein the system (1) further comprises a controller (7) andat least one sensor (70, 72, 74, 76) in communication with saidcontroller, wherein the system (1) comprises at least one temperaturesensor (70) arranged upstream of the pressure regulator (4) anddownstream of the heat exchanger section (2A), wherein the controller(7) is configured to control the first pressure regulator (4) based onthe measured value of the at least one temperature sensor (70) tocontrol the state of the two-phase process medium; the system (1)comprises at least one filling sensor (72) arranged in the vessel (5)and/or at least one flow sensor (76) arranged downstream of the pressureregulator (4) for measuring a mass flow of a liquid phase of the processmedium to the load, wherein the controller (7) is configured to controlthe pressure regulator (4) to control the mass flow based on themeasured value of the at least one filling sensor (72) and/or the atleast one flow sensor (76); and/or the system (1) comprises at least onepressure sensor (74) arranged in communication with the vessel (5) and acompressor system coupled to the outlet (36) of the heat exchanger (3),wherein the controller (7) is configured to control the pressure in thevessel (5) by controlling the compressor system based on the measuredvalue of the at least one pressure sensor (74).
 9. Cryogenicrefrigeration system (1) according to claim 8, wherein the system (1)further comprises a control valve (20) for controlling the mass flow ofthe supply flow (10), which is in communication with the controller (7)and is arranged in parallel to and upstream of the first pressureregulator (4), wherein the controller (7) is configured to control themass flow of the supply flow (10) via the control valve (20) based onthe measured value of the at least one temperature sensor (70), fillingsensor (72) and/or flow sensor (76).
 10. Cryogenic refrigeration system(1) according to claim 8, wherein the system (1) comprises at least onewarm-end temperature sensor (70) in communication with the conduit (2)and the outlet (36) of the heat exchanger (3) at the warm end (32) ofthe heat exchanger (3), wherein the controller (7) is configured toadjust the evaporative gas flow based on a temperature differencemeasured by the at least one warm-end temperature sensor (70) bycontrolling the pressure regulator (4).
 11. Cryogenic refrigerationsystem (1) according to claim 1, wherein the heat exchanger (3) isconfigured as a plurality of heat exchanging modules (3A, 3B, 3C), whichare arranged in parallel and/or in series to the conduit (2), whereinpreferably a second pressure regulator (4B) in fluid communication withthe conduit (2) is arranged between each serially arranged heatexchanging module (3A, 3C).
 12. Method for providing a cryogenicrefrigeration in a cryogenic refrigeration system (1), the methodcomprising: providing a supply flow (10) of a process medium in aconduit; cooling the supply flow in a counter flow heat exchanger (3);reducing the pressure of the supply flow (10) by means of a pressureregulator (4); and receiving the supply flow (10) in a vessel (5),wherein an evaporated gas flow from the process medium is used by theheat exchanger (3) to cool the supply flow (10), wherein the cooling ofthe supply flow is provided free of any evaporating liquid phase. 13.Method according to claim 12, characterized in that a temperature factorof the evaporated gas at a warm end (32) of the heat exchanger (3)relative to the process medium of the supply flow (10) at the warm end(32) of the heat exchanger (3) is provided by the heat exchanger, whichis larger than 0.9, preferably larger than 0.98, during normal operationof the cryogenic refrigeration system (1); and/or a temperature of theevaporated gas is matched to a temperature of the process medium at awarm end (32) of the heat exchanger (3) during normal operation of thecryogenic refrigeration system (1) provided by an NTU configuration ofthe heat exchanger (3).
 14. Method according to claim 12 or 13, whereinthe supply flow (10) comprises a pressurized liquid, preferably liquidhelium, wherein reducing the pressure of the supply flow (10) by thepressure regulator (4) provides a two-phase process medium (13) flowdownstream of the pressure regulator (4) and wherein the evaporated gasin the vessel is provided at sub-atmospheric pressure, wherein thecooling of the supply flow (10) preferably provides the process mediumbetween the lambda point and the saturation temperature downstream ofthe heat exchanger section (2A) of the conduit (2), and wherein thevessel (5) preferably collects the liquid phase (15) of the processmedium to refrigerate a thermally coupled load (6) or a load (6)disposed in the liquid phase (15) of the process medium in the vessel(5), to provide an isothermal load (6).
 15. Method according to claim12, wherein the cooling of the supply flow (10) occurs in series or inparallel by means of a plurality of heat exchanger modules (3A, 3B, 3C)arranged in series or in parallel, wherein preferably the pressure ofthe supply flow (10) is reduced between each serially arranged heatexchanger module (3A, 3C) by means of a second pressure regulator (4B).